Rotary shaft control apparatus

ABSTRACT

Apparatus for controlling a rotary shaft relative to a reference includes a strain wave drive having a specially shaped deformable ring gear for minimum stress. The wave generator part of the drive is normally caused to rotate with the input and output shafts connected to the drive&#39;s ring gears so that the two shafts are effectively keyed together. A tractive-type clutch assembly permits the wave generator to be rotated in one direction or the other relative to the deformable gear shaft so that the phase angle of the two shafts can be adjusted at a slow rate of speed dependent on the gear ratio of the drive. Also a tractive-type planetary assembly may be coupled between the clutch assembly and the wave generator so that the phase change may be accomplished at a plurality of speeds.

RELATED APPLICATIONS

This is a division of application Ser. No. 268,840, filed June 1, 1981,now U.S. Pat. No. 4,382,391, which is a division of application Ser. No.965,236, filed Dec. 1, 1978, now U.S. Pat. No. 4,286,476.

BACKGROUND OF THE INVENTION

Many manufacturing processes involve the close coordination of a seriesof steps that are synchronized cyclically in automatic machinery. Thepower and motion required for the operation is usually introduced intothe machine by means of sprockets or gears affixed to certain shafts bymeans of shaft keys, and in a like manner the energy and appropriatemotions are carried throughout the machine by other shafts and otherkeyed sprockets (gears) to perform the specific steps required. Due tovarious material behaviors and eccentricities characteristic ofmachinery, the desired timing relationships between shafts is not alwaysforthcoming. Means must be available to adjust these various shaftingtiming relationships dynamically while the machine is operating.

Since sprockets and gears are keyed to their shafts, no dynamicadjustment is normally possible between a sprocket and its own shaft andany correction necessary must be performed by an external mechanismacting between any two shafts that are required to be timed, rather thanbetween a particular sprocket and its own shaft. Power then, istransmitted through such an external mechanism from one machine shaft toanother, but in so doing, a change in phase angle is caused to existbetween those two machine shafts. Usually this external phase changingmechanism is a controllable differential that can advance or retard thephase relationship. Depending on the design this control can be manualor remote, or both. In most designs the motion passing through iscontinuous, with internal gears rotating constantly with respect to oneanother.

Such an external phasing mechanism is wall or floor mounted, having aninput and output shaft. Power from the first machinery shaft istransmitted by a chain to the input of the differential and thence byanother chain from the differential output to the machinery shaft numbertwo. A multiplicity of shafts require almost a jungle of chains thatflow to and from a multiplicity of wall or floor mounted differentials.In addition, each differential requires its own actuating motor or powersource to change the phase angle.

For example, many web tension control systems use draw rollers driven bya controllable differential which controls the tension in a webconducted to a web-consuming machine such as a press. The differentialis controlled by the output of a tension sensing device such as amovable dancer roller around which the web is trained, with theexcursions of the dancer from a fixed reference position being detectedby limit switches or the like.

Another application in which the precise timing control of a shaft isessential is the coordination of two or more conveyors. For example, ina given production line it may be necessary to transfer a succession ofarticles advancing on a horizontal conveyor onto a succession ofvertically advancing shelves on a vertical conveyor. Here again, thespeeds of the two conveyors may be synchronized by driving one of theconveyors from the main drive by way of a controllable differential. Thedifferential is set to advance or retard the one conveyor relative tothe other as needed to assure that the vertical conveyor is in thecorrect timing position to receive each article leaving the horizontalconveyor.

The maintenance of accurate control over a rotary shaft relative to itsdriver is especially important in a printing press wherein web isadvanced past one or more printing cylinders which print the printedmatter in one or more colors. Not only must the motions of the printingcylinders be controlled precisely relative to the advancing web toposition the printed matter at the proper location on the web, but alsothe motions of the cylinders must be synchronized with each other sothat the print from each cylinder is in exact registration on the web.

Finally, when controlledly driving shafts or rollers from a main driveshaft, it may be desirable to adjust the phase angle of the rollerrelative to the main shaft in a forward or reverse direction forpurposes of registration or synchronization. During normal operation, itis desirable that the phase change take place quite slowly to avoidovershoot and for optimum accuracy. On the other hand, during initialset-up at the start of a run when the draw rolls are completely out ofphase with the drive shaft or in times of emergency, it is desirablethat the phase correction take place more quickly to save time and toavoid wastage. All other shaft control apparatus of this general typerequire external power sources in order to accomplish this and so havethe same disadvantages described above.

A more convenient and satisfactory solution to the aforementionedcontrol problems is the provision of a mechanism that mounts on a shaft,acting as an adjustable key, capable of changing the rotational positionof a sprocket relative to its own shaft, dynamically and in eitherdirection, drawing the power required for this function from the motionof the sprocket or its shaft without an ancillary power source.

Such a mechanism, then, would be useful in any field requiring machinetiming. Such fields, to mention a few, include printing, packaging,collating, machinery timing, boxing, synchronizing, automatic assembly,tension control, and phase shifting.

SUMMARY OF THE INVENTION

Accordingly the present invention aims to provide improved apparatus forcontrolling the angular or phase relationship between a rotary shaft andanother element.

Another object of the invention is to provide improved shaft timing andregistering apparatus.

Another object is to provide such shaft control apparatus which can haveexceptionally fine control capabilities.

A further object is to provide such apparatus whose reduction ratio canbe changed easily by a simple interchanging of a part or parts.

A further object of the invention is to provide apparatus of this typewhich permits the speed and/or phase adjustment of a rotary shaft whilethe shaft is running.

Yet another object of the invention is to provide apparatus of thisgeneral type which is completely mechanical, requiring no additionalpower source to effect the shaft speed or phase change.

Another object is to provide apparatus for advancing and retarding aharmonic strain wave drive generator dynamically relative to its shaft.

Still another object of the invention is to provide a compact,shaft-hung apparatus which functions as a variable shaft key to adjustthe phase angle of a shaft and its driving sprocket.

Another object is to provide such apparatus having essentially zerobacklash.

A further object of the invention is to provide apparatus of this typewhich can effect shaft speed or phase angle changes at a plurality ofdifferent speeds.

Another object of the invention is to provide apparatus of this typewhose internal parts for the most part move relative to one another onlywhen a speed or phase change is actually being effected.

Yet another object of the invention is to provide rotary shaft controlapparatus which provides exceptionally high torque coupling (i.e. 300inch pounds) between its input and output shafts.

A further object of the invention is to provide rotary shaft controlapparatus which is relatively easy and inexpensive to manufacture andassemble and is therefore relatively inexpensive.

Another object is to provide an improved strain wave drive for use inapparatus of this general type.

A further object is to provide a drive of this type whose deformablegear suffers a minimum amount of stress in operation.

Still another object is to provide an improved drive of this type whichrequires a minimum of axial load in order to operate.

A further object is to provide an improved strain wave drive which isrelatively easy and inexpensive to manufacture.

Another object is to provide an improved clutch or brake assembly foruse particularly in a shaft control apparatus to effect shaft motion inboth directions, especially as applied to the wave generator in strainwave designs.

Yet another object is to provide an improved planetary assembly for useparticularly in a shaft control apparatus to obtain especially high gearreductions, especially as applied to control the wave generator ofstrain wave drives.

Other objects will, in part, be obvious and will, in part, appearhereinafter.

The invention accordingly comprises the features of construction,combination of elements and arrangement of parts which will beexemplified in the following detailed description, and the scope of theinvention will be indicated in the claims.

In general, my shaft control apparatus can be used in almost anyapplication involving control of the speed or phase angle of a rotaryshaft relative to a speed or phase reference. The reference may be asecond driven shaft or an event in time or position. The presentinvention has particular application as a variable shaft key for varyingthe phase relationship between a sprocket and its shaft. Accordingly,the invention will be described in that context. It should beunderstood, however, that it has equal functionality in regulating,synchronizing, metering and other similar applications involving themaintenance of control of a rotary shaft relative to a reference.

The control apparatus includes an inner shaft and a concentric tubularouter shaft. The inner shaft is usually arranged to be supported by, andfixed for rotation with, the shaft being controlled, while a sprocketdriven by a timing chain or the like is fixed for rotation with theouter shaft. However, the roles of the inner and outer shafts could justas well be reversed. Thus the entire apparatus is mounted on andsupported by the shaft being controlled.

The apparatus inner and outer shafts are speed and phase coupled bymeans of a strain wave drive of the general type disclosed in U.S. Pat.Nos. 2,906,143; 2,943,508; 3,001,890; 3,117,763; and my U.S. Pat. No.3,187,605. A drive of that type comprises a pair of coaxial annularinternal and external ring or conical gears, one of which is elasticallydeformable, the other being rigid. A strain wave generator engages thedeformable gear to deflect a working surface thereof into engagementwith a working surface of the rigid gear at a plurality ofcircumferentially spaced positions interspaced by non-mating positions.In effect then, the two gears are keyed together at those engagingpositions.

In the present apparatus, the outer ring gear is formed in a surface ofthe tubular outer shaft and the inner gear is formed in a flexibleresilient annular surface of revolution positioned concentrically insidethe outer shaft and projecting axially from the end of the inner shaft.Preferably, although not necessarily, the inner gear has a specialconical-type shape to be described in detail later which minimizesstresses on the gear and on other parts of the apparatus. The drivenstrain wave generator is rotatively mounted on the inner shaft insidethe inner ring or conical gear.

The general operation of a harmonic or strain wave drive is well known.Usually the deformable ring is fixed against rotation and the strainwave generator is rotated at a selected speed. This produces a strainwave which rotates about the deformable gear and causes the rigid gearto rotate in the same direction as the wave generator at a reduced speedrelative to the wave generator dependent upon the number of teeth in thetwo ring gears. For example, if the rigid outer ring has 160 teeth andthe inner deformable ring has 150 teeth, the drive will provide a gearreduction ratio of 80 to 1.

In the present apparatus, however, the rotary input is applied, not tothe wave generator, but to the rigid outer gear by way of the sprocketand the torque output is taken from the deformable inner gear connectedto the inner shaft rotatively fixed to the shaft being controlled. Thewave generator always rotates with the deformable gear connected to theoutput shaft as long as it is not being controlled externally. With thisarrangement, as long as the strain wave generator rotates with theoutput shaft, the outer gear and its sprocket remain rotatively fixed tothe shaft. Thus, in effect, the strain wave drive functions as a splinewhich rotatively keys the sprocket to the shaft so that the two rotatein phase in the same direction.

Further, provision is made in the present apparatus for controlledlyrotating the wave generator in one direction or the other relative tothe deformable gear secured to the output shaft. This causes relativerotation between the inner and outer ring or conical gears. Accordingly,it permits the driven shaft to be advanced or retarded in phase relativeto the sprocket providing, in effect, a variable angle spline betweenthe sprocket and the driven shaft.

The relative motion between the wave generator and the deformable gearis provided by a special traction-type clutch assembly to be describedlater. Suffice it to say at this point that the clutch assembly can becontrolled to stop the rotation of the wave generator relative to thedeformable gear so that the wave generator has, in effect, a velocity of-V relative to that gear where V is the velocity of that gear. Thiscauses the output shaft connected to the deformable gear to be advancedin phase relative to the sprocket at a fraction of that speed dependentupon the number of teeth in the two gears i.e., 1/80V in the aboveexample.

The clutch assembly can also be controlled to cause the wave generatorto rotate at twice the speed of the sprocket in the same directionthereby causing the output shaft to be retarded in phase relative to thedeformable gear at the reduced speed characteristic of the particularstrain wave drive e.g. 1/80V. Using the clutch assembly, then, while theapparatus is running, the output shaft can be advanced or retarded inphase relative to the sprocket at a rate dependent upon the speedreduction ratio of the particular strain wave drive. For example, thephase of a draw roller connected to the output shaft can be advanced orretarded relative to a main drive coupled to the sprocket to synchronizethe phase of the roller with those of other shafts and rollers operatingoff the main drive.

By the same token, the apparatus can be controlled automatically by theoccurrence of remote events. For example, its clutch assembly can bearranged to respond to excessive excursions of a tension sensing dancerso as to advance or retard a web draw roller relative to the line speedto maintain substantially constant tension in a web.

In some applications it is desirable to make the phase change betweenthe output shaft and the sprocket at a plurality of rates. For example,when bringing a roller into registration with a drive shaft, it may bedesirable to use a relatively fast rate of correction until the rolleris nearly in register and then switch to a slower or finer rate ofcorrection to achieve actual registration. The present apparatusprovides for such plural speed phase correction in that a special speedreducing traction planetary assembly to be detailed later may beselectively coupled between the clutch assembly and the strain wavedrive. This permits the relative angular velocity of the strain wavegenerator to be made a fraction of the output shaft speed, e.g. 19/20Vor smaller.

Accordingly when the planetary assembly is active, the output shaft canbe advanced or retarded relative to the sprocket at a very small rateequal to the product of the speed reduction provided by the planetaryassembly and the speed reduction provided by the strain wave generator,i.e. 1/20×1/80 or 1/1600V. This is equivalent to about 45° per minutewith the output shaft running at 200 rpm. Using the planetary assembly,speed reduction ratios as high as 40,000 to 1 or higher can be obtained.On the other hand, when a faster rate of correction is desired, theplanetary assembly can be decoupled from the system permitting the phasecorrection to take place at the 1/80th relative speed ratio provided bythe strain wave drive alone, which is equivalent to about 900° perminute with the output shaft rotating at the 200 rpm speed.

During normal operation, when the sprocket is in the correct phaserelationship with respect to the output shaft and the clutch assembly isdisengaged so that the strain wave generator is not subjected toexternal control, all of the components of the strain wave driveincluding the inner and outer ring gears and the strain wave generatormove in unison. Consequently, all torque is coupled directly between theinput and output shafts; the strain wave generator does not bear any ofthis load. Furthermore, during such operation, substantially all of therotary components of the apparatus move in unison so that there issubstantially no parts-wearing relative motion between them as in thecase with the prior speed differentials described above.

Torque is indeed coupled through the wave generator when it is necessaryto effect a phase change between the input and output shafts. However,it is only coupled in one direction, i.e. from the wave generator to thering gears and not vice versa. Consequently, even during a phase change,the strain wave generator is not subjected to any appreciable part ofthe load torque. In the case where the strain wave drive ratio is 80:1,then, the wave generator has to transmit only 1/80th of the line torqueduring phase changing.

Finally since the apparatus is shaft-mounted and makes the phasecorrections between the input and output shafts without any externalmotive means, it is relatively inexpensive and occupies a very smallamount of space so that it can easily be retrofit on existing machinerywhere available space is at a premium.

BRIEF DESCRIPTION OF THE DRAWINGS

For a fuller understanding of the nature and objects of the invention,reference should be had to the following detailed description taken inconnection with accompanying drawings in which:

FIG. 1 is a perspective view with parts broken away showning the rotaryshaft control apparatus of this invention arranged to couple torquebetween the sprocket and a roller;

FIG. 2 is a sectional view on a larger scale of the FIG. 1 apparatus;

FIG. 3 is a sectional view along line 3--3 of FIG. 2;

FIG. 4 is an exploded perspective view showing various components of theFIG. 1 apparatus in greater detail;

FIG. 5 is a sectional view with parts cut away along line 5--5 of FIG. 2on a smaller scale showing the clutch assembly of the FIG. 1 apparatuswith the clutch disengaged;

FIG. 6 is a view similar to FIG. 5 with certain elements omitted forclarity, showing the clutch engaged;

FIG. 7 is a sectional view along line 7--7 of FIG. 2 showing the speedshift key in one position;

FIG. 8 is a sectional view along line 8--8 of FIG. 7 with parts cut awayfor clarity showing the shift key in its other position;

FIG. 9 is a fragmentary perspective view showing the shift key camportion of the apparatus in greater detail, and

FIG. 10 is a similar view illustrating the operation of the planetaryportion of the apparatus.

DESCRIPTION OF AN ILLUSTRATIVE EMBODIMENT

Refer now to FIG. 1 of the drawings which shows my control apparatusindicated generally at 10 arranged to control a rotary member 12. Member12 may be a draw roller, a rotary cutter, a printing cylinder or thelike. For purposes of this description, we will assume member 12 is adraw roller. Roller 12 has a shaft 14, the opposite ends of which arerotatively mounted in bearing units 16, only one of which is shown, theunits 16, in turn, being anchored to a machine frame 18. The illustratedend of shaft 14 extends beyond fixture 16 and is fitted with the usualkey 22. The apparatus 10 is mounted directly on the projecting end ofshaft 14, although in FIG. 1 it is shown separated from the shaft inorder to show the interconnection between the two more clearly.

The control apparatus comprises a strain wave drive section indicatedgenerally at 24, a speed reducing planetary section shown generally at26, a clutch section indicated at 28 and a control section showngenerally at 30.

The drive section 24 includes a tubular inner shaft 32 and a concentrictubular outer shaft 34 which is rotatable relative to shaft 32. Toconnect the apparatus to the roller 12, the projecting end of rollershaft 14 is inserted into shaft 32, the latter shaft being provided withan internal keyway 38 to accommodate the shaft key 22. Thus when theapparatus is mounted on the shaft 14, the shaft 14 and shaft 32 arerotatively coupled together.

Secured to the end of the outer shaft 34 is a sprocket 42 which isprovided with a central opening 44 to loosely accommodate shaft 14. Thesprocket is secured to the shaft 34 by means of bolts 46 which areturned down into threaded openings 48 in the end of shaft 34. Thus thesprocket 42 and shaft 34 are rotatively coupled together. Strengtheningdowels may be inserted through the sprocket and shaft if necessary. Thesprocket is connected via a timing chain 52 to a main drive shaft orother comparable rotary motive source (not shown). Thus the apparatus 10is suspended from and entirely supported by the roller shaft 14.

When the sprocket 42 is rotated by the timing chain 52, the apparatusdrive section 24 couples torque directly to the roller shaft 14 so thatthe rotation of the two is identical in both direction and phase. Thusfor all intents and purposes, the apparatus functions as though thesprocket 42 is keyed directly to the shaft 14 rather than by way of theapparatus 10.

In many cases, however, it is desirable to change the phase anglebetween roller 12 and the sprocket 42 for one reason or another such asto bring roller 12 into registration or to synchronize it in phase withanother roller driven by the same main drive shaft.

Assume, for example, that it is desired to rotate the roller 12 relativeto the sprocket so as to bring the arrow A on the roller intoregistration with the arrow B on the sprocket. Assume further that thesprocket is rotated in the clockwise direction as indicated by the arrowS in FIG. 1. By means of the switch control buttons in the controlsection 30 this can be accomplished quite easily. More particularly,section 30 includes an ADVANCE control button 53 which when depressedcauses the clutch section 28 to rotate the output shafts 14 and 32 at afaster rate than the input shaft 34 and its sprocket 42. Consequentlythe roller 12 is advanced in phase relative to the sprocket. As soon asthe two arrows A and B are aligned, the button 53 may be released.Whereupon, the input and output shafts again rotate in unison with theapparatus 10 functioning effectively as a fixed spline between them.

Section 30 also includes a RETARD switch control button 54 which whendepressed causes the clutch section 28 to rotate the output shaft 32 ata slower rate than the input shaft 34. Necessarily, then, this causesthe roller 12 to be retarded in phase relative to sprocket 14. Usingthat button 54, the arrow A on roller 12 can be brought intoregistration with an arrow C on sprocket 42.

Control section 30 also provides for achieving the aforesaid relativerotation between the roller and sprocket at a plurality of differentspeeds. More particularly, section 30 includes a SPEED CONTROL switchbutton 56. When the button 56 is in its depressed position, the clutchsection 28 acts directly on the drive section 24 so that the relativemovement between the roller and sprocket, which occurs when eitherbutton 53 or 54 is depressed, proceeds at a relatively rapid rate. Onthe other hand, when the button 56 is in its raised position, followedby depression of button 53 or 54, the planetary section 26 is coupledbetween the clutch section 28 and the drive section 24 with the resultthat the phase correction between the roller and the sprocket occurs ata much closer rate.

Thus by properly manipulating the controls in section 30, the apparatuscan be made to vary the phase angle between the roller 12 and thesprocket 42 and its main drive shaft in both directions and at twodifferent speeds.

During normal operation of the apparatus, when the sprocket 42 is in thecorrect angular relationship with the roller and the two are rotated inunison, there is no net torque applied to sections 28 and 30 at the freeend of the apparatus 10. However when the apparatus is functioning tochange the relative speed of the sprocket and roller, there is arelatively small reaction torque imparted to those sections which tendsto cause them to rotate about the axis of shaft 14. To prevent this, atorque arm 58 is connected at one end to section 30, its opposite endbeing captured in some convenient manner so that is cannot swing ineither direction about the shaft 14 axis. However, it should beunderstood that the torque arm 58 does not play any part in supportingthe shaft-mounted apparatus 10 and the reaction torque is quite smalldue to the high, e.g. 80:1, gear reduction provided by the drive section24.

It should be emphasized also that the apparatus depicted in FIG. 1 iscomplete in that it uses the driving power that passes between thesprocket 42 and the roller 12 to change the relative phase or speed ofthe sprocket and roller. Accordingly, no auxiliary motor is required forthat purpose. Consequently, the entire apparatus can be mounted at theend of shaft 14. For this reason and because it is quite small andcompact, the control apparatus can be retrofitted on many existinginstallations where such shaft control is desired.

Referring now to FIGS. 2 and 3 of the drawings, the strain wave drivesection 24 of the apparatus comprises an inner, elastically deformablering gear 62 which is secured to the inner end of tubular shaft 32extending out coaxially therefrom. The ring gear 62 is an annularsurface of revolution which may be cylindrical as described in U.S. Pat.No. 2,906,143 or conical, spherical or other surface of revolution asdepicted in my U.S. Pat. No. 3,187,605.

Most preferably, however, it is in the form of an annular surface ofrevolution having a so-called "hour-glass" shape as viewed in medialsection. By this, I mean that the gear 62 has an end section 62a whichis conical with a relatively large maximum diameter. Section 62a tapersdown to an intermediate section 62b having a much smaller diameter. Thatsection, in turn, flares radially outwardly along a generally circulararc to the cylindrical section 62c which is attached to shaft 32 andwhich has a larger diameter than section 62b.

The end of the gear section 62c is soldered, brazed or welded to thesleeve by a circumferential bead 64. Also the inner end section 32a ofshaft 32 is tapered to provide a clearance space between the shaft andgear 62 to permit the latter to deflect during operation of the strainwave drive section 24. The actual teeth 66 (FIG. 3) of the gear 62 areformed only in the outside surface of the conical gear section 62a.These teeth 66 are spaced equidistant from one another all aroundsection 62a.

The strain wave drive section 24 also includes a rigid outer ring gear68 which is formed adjacent the inner end of the outer shaft 34. Gear 68is preferably a conical annular surface of revolution having more orless the same pitch as the deformable ring gear section 62a. It also hasa set of gear teeth 72 which have the same spacing and orientation asthe inner ring gear teeth 66. However, as is customary with strain wavedrives of this general type, the outer ring gear is slightly larger indiameter than the inner ring gear permitting it to have a few moreteeth. The actual number of teeth in the two gears depends upon thespeed reduction desired to be provided by the drive section 24. In theembodiment of this invention being described, the outer ring gear has160 teeth while the inner ring gear has 158 teeth. Consequently when thedrive is operative during a phase angle correction, it provides an 80 to1 speed reduction. The specific relationship between the number of gearteeth and the speed reduction provided by such a drive is well known andwill not be detailed here.

Still referring to FIG. 2, snugly received inside the inner end oftubular shaft 32 and secured thereto is a sleeve 73 in which one end ofa shaft 74 is keyed. The shaft 74 is locked to shaft 32 by a pin 76extending through diametrically aligned openings in shaft 32, sleeve 72and shaft 74. Thus the shaft 74, being an extension of output shaft 32,is also an output shaft and it is fairly long so that its opposite endprojects beyond the apparatus clutch section 28. A washer 78 having aninner diameter somewhat smaller than the inner diameter of shaft 34 ispositioned on the end of the shaft inboard of sprocket 42 as shown inFIG. 2. The washer is rotatively engaged by a split ring 82 seated in anexternal circular groove 84 in shaft 32 situated just beyond the washer.The washer is held in place by the bolts 46 that secure the sprocket tothe end of shaft 34. If desired, a spacer 85 having a reduced diameterend 85a may be included between the sprocket and washer 78 to concealthe end of the inner shaft 32. However, in that event there should be aclearance space between shaft 32 and spacer end 85a.

Still referring to FIGS. 2 and 3, the final component of the strain wavedrive section 24 is a strain wave generator shown generally at 86 inFIG. 2. Generator 86 includes a rotary cam 88 having a central opening92 for rotatively receiving shaft 74. Formed in cam 88 is a cam surfacein the form of an elliptical bearing race 94 facing the ring gearsection 62a.

Riding in race 94 is an annular set of bearings 96 distributed in aflexible outer bearing race 98 having a conical outer surface 98aarranged to fit flush against the inside surface of the gear section62a. As best seen in FIG. 2, the race 94 is axially offset in asinusoidal manner so that at, say, 0° and 180° its surface is relativelyclose to gear section 62a whereas at 90° and 270°, its surface isfurther away from that gear section. Actually, the displacement of therace varies as a function of (1-cos θ) in a direction about 23.5° fromthe cam axis.

When the cam 88 is loaded axially against gear 62, the flexible bearingrace 98 is pressed against the gear at diametrically opposite locationsthereon corresponding to the positions of the diametrically oppositeraised portions of the race 94, i.e. near 0° and 180° in FIG. 2. Thisaxial force deforms gear section 62a from its natural round (incross-section) shape into an oval or elliptical shape so that the gearteeth 66 adjacent the major axis of the ellipse are pressed intoengagement and mesh with the opposite teeth 72 in the outer ring gear 68as best seen in FIG. 3. On the other hand, the gear teeth 66 adjacentthe minor axes of the ellipse, i.e. at 90° and 270°, are drawn away fromthe corresponding teeth in the outer ring gear sufficiently to clearthem also as shown in FIG. 3, the clearance space being exaggerated forclarity.

If now, the cam 88 is rotated on shaft 74 relative to inner gear 62, thediametrically opposite points of toothed engagement between the innerand outer ring gears 62 and 68 will rotate commensurately. As is wellknown with strain wave drives of this type, if the inner gear is fixed,such rotation causes relative rotary motion between the outer ring gear68 and the wave generator 86 in the same direction at a greatly reducedspeed dependent upon the number of teeth in the two ring gears i.e. an80:1 speed reduction in this example. On the other hand, if the outergear is fixed, the inner gear rotates in the opposite direction relativeto the wave generator at the same reduced rate.

The illustrated gear 62 is an improvement on the one described in myaforementioned U.S. Pat. No. 3,187,605. That patent describes adeformable element for a strain wave drive consisting of anon-cylindrical body of revolution which is adapted for deflection withthree degrees of freedom. Among the illustrative gear embodiments inthat patent is a conical body of revolution which produces significantadvantages over the then known prior art in terms of ease of manufactureand flexibility in the choice of gear characteristics. While the stressimposed upon that deformable gear in use is significantly less than thatencountered by the cylindrical gears of the type described in theaforementioned U.S. Pat. No. 2,906,143, the stress is still higher thanit should be.

This is because in order for the gear to have a high torque couplingcapability, e.g. 300 inch pounds, its wall thickness must be relativelylarge. This of course makes the gear quite stiff. Consequently, the wavegenerator which deflects that gear against the rigid outer gear duringoperation of the drive must be highly loaded axially against teh gear.Indeed axial forces as large as 400 pounds are required in order toenable the drive to operate properly. Such a high loading force canimpose stresses on the deformable gear as high as 78,000 psi or more. Inorder to be able to withstand these stresses, the gear must be made ofexpensive material such as 52100 steel which is difficult and expensiveto machine. Even then, the operating life of the gear is not as long asit might be. In addition, however, that high axial load is also impartedto the other components of the drive such as bearings, rings, cams, etc.Consequently, those parts as well must be fabricated using expensivematerials which must then be machined to tolerance.

While analyzing the operation of these deformable gears, I found that anon-cylindrical body of revolution such as a cone as the deformable gearin a strain wave drive functions more or less like a spring. In otherwords, there is a substantially linear relationship between thedeflection of the toothed rim of the gear and the axial load applied tothe wave generator deflecting the gear, which load is proportional tothe axial displacement of the wave generator, i.e. deflection equals theproduct of a constant and axial displacement which is the equation for asimple spring. I found also that due to the angulation of a conicalsurface of revolution, its absolute slope decreases as the geardisengages from the rigid rear teeth during operation of the drive.Consequently, I concluded that it is not necessary to deflect the gearby an amount equal to twice the tooth height as is usually done in suchdrives. Rather the drive should operate properly if the deformable gearis deflected by a lesser amount equal to twice the tooth height minusthe angulation factor. That factor can be particularly large in the caseof a cone.

Thus for example, a cone angle of 22.5° as measured between the coneaxis and one side, a 60° tooth angle and 160 teeth for a chosen diameterwill yield a tooth measuring about 0.032 inch. This normally implies(using the standard deflection of twice the tooth height) a differentialbetween the major and minor axes of the deformable gear of 2 times 0.032or 0.064 inch. However, my experiments have shown that in the case of acone, this difference need only be about 0.031 inch, or about half thatamount.

Consequently, a relatively small axial loading force on the drive, e.g.40 pounds, should suffice provided the gear has a sufficiently lowspring constant.

The spring constant in the case of such bodies of revolution is afunction of several parameters such as length, diameter, wall thickness,type of material and also the shape of the body between the portionthereof carrying the teeth 66 and its connection to the torque take-offmeans, i.e. shaft 32. Of these factors, I decided that the gear shapewould have a particularly significant effect on the spring constant ofthe gear.

A gear composed of a non-cylindrical body of revolution such as thepresent one naturally tends to angulate in the manner of a system ofcantilevered beams connecting the ring of teeth formed thereon at thepolar axis of rotation of the doby as described in my patent. However,the connection of a torque take-off means such as a sleeve or shaft tosuch a gear in the usual way at a location spaced from the polar axisupsets the normal conical angulation of the gear and thereby causes theteeth to angulate about the boundary between the torque take-off meansand the gear. This produces unnatural stresses in the gear. Moreimportantly it renders the gear stiffer than it would be if it couldangulate naturally.

In a cone, for example, portions of the body tend to angulate about theapex of the cone lying on the polar axis. Therefore the formation of atorque take-off sleeve on or integral with the cone and which is locatedan appreciable distance from that apex toward the teeth prevents thetoothed portion of the cone from angulating properly about the apex.Rather, the cone is constrained to angulate about the circular boundarybetween the cone and the sleeve thereby rendering the gear stiffer andsubjecting it to unwanted stress. In attempting to avoid this problem, Iconcluded that since a conical body naturally angulates about its apex,the torque take off means should be connected to the conical portion ofthe gear at a maximum diameter such that the conical action is stillmaintained while at the same time providing a sufficient polar moment ofinertia in the gear to permit taking off a useful torque.

In this connection, it should be understood that motions do exist in themedial plane P₁ (FIG. 2) that passes through the minimum diameter of thegear. These motions are both radial and axial and are different aroundthe polar axis of the gear. Any further means secured to that plane totake off torque can inhibit those motions and increase the stress in thegear.

I concluded, therefore, that the portion of the gear beyond that minimumdiameter plane should increase in diameter with a smooth transitionalong more or less mirror image circular arcs A₁ and A₂ to a new medialplane p₂ in which such radial and axial motions are not present,although they do exist between the two planes. At the latter plane P₂the first contact with and connection to the torque take-off means,i.e., shaft 32, is made.

In other words, the connection should be made along a circular boundaryat P₂ positioned radially outward along mirror image spherical segmentsfrom the smaller diameter end of the conical section at P₁ as best seenin FIG. 2. In this way, the securement of the torque take-off means 32to the gear 62 has minimal effect on the natural angulation of thetoothed conical section 62a.

For purposes of the present apparatus 10, the gear 62 illustrated inFIG. 2 provides the desired toothed engagement with the outer ring gearwith an axial load on the wave generator of only about 40 pounds.Therefore the stress imparted to the gear when the drive is operative isrelatively low, i.e. about 28,000 psi. Consequently, the gear can bedie-formed in a few steps from an ordinary inexpensive cold-rolled steeltube. Further, the other axially loaded components of the apparatus canlikewise be cast or otherwise formed easily from inexpensive materialssuch as powdered metal which cast parts are quite able to withstand suchrelatively low stress.

As best seen in FIGS. 2 and 8, seated on the inner end of shaft 34 is acircular collar 102 having a reduced diameter neck 102a which fitsinside shaft 34 and rotatively engages a shoulder 88a of the rotary cam88. The collar 102 is keyed at 103 for rotation with the shaft. The cam88 has a reduced diameter neck 88b which projects along shaft 74 more orless to the end of the collar 102. The exposed end of collar 102 isbeveled and formed with teeth so that it constitutes a female bevel gear104 (FIG. 8) which actually constitutes part of the speed reducingplanetary section 26 is to be described presently.

Turning now to FIGS. 2 and 4, the planetary section 26 includes aso-called reduction tube 110 which is a tubular member rotativelysupported on shaft 74. Tube 110 is formed with a radial flange 110ahaving a circumferential array of openings 112 which form cages for acorresponding array of ball bearings 114. Tube 110 is rotativelyreceived on shaft 74 at a location spaced from collar 102. Rotativelyseated on a reduced diameter neck 110b on tube 110 is a circular ring orplate 116. The side face of ring 116 facing the tube flange 110a isformed with a circular race 118 (FIG. 2) for receiving the bearings 114.The opposite face of the ring is formed with a similar circular race122.

Rotatively mounted on tube 110 to the left of ring 116 is a so-calleddrive tube 124. Tube 124 is similar to tube 110 in that it has a radialflange 124a formed with a circular array of openings 126 for caging anarray of ball bearings 128. These bearings are arranged to ride in therace 122 in ring 116.

Rotatively mounted on the drive tube 124 is a second ring 132. The sideof the ring facing tube 124 is formed with a circular bearing race 134(FIG. 2). Also received on drive tube 124 to the left of ring 132 is acircular thrust bearing 136 illustrated in FIG. 2. That bearing is notshown in FIG. 4 for reasons of drawing clarity. Finally, positioned ontube 124 beyond the thrust bearing is an annular reduction disk showngenerally at 138 whose function will be described later. Suffice it tosay at this point that the disk 138 has an inner opening 140 engagingaround tube 124. Also a circumferential array of axial keys 142 areformed in the wall of opening 140 which interfit with correspondinglyshaped and located keyways 144 in the end of the drive tube 124 so thatthe drive tube and disk 138 rotate together.

As best seen in FIG. 2, when the drive tube 124 with all of itsencircling components is engaged on tube 110, the end of the tube 110projects somewhat beyond the end of tube 124. That projecting end isprovided with a pair of diametrically opposite axial keyways 152 (FIG.4) which are arranged to receive correspondingly shaped and located keys154 formed in the inside wall of the cam neck 88b. Also, a thrustbearing 156 (FIG. 2) is engaged around the cam neck which reacts betweenthe shoulder 102a of ring 102 and the reduction disk 138.

Thus when the planetary section 126 is coupled to the drive section 24and more specifically when the reduction tube 110 is keyed to the camneck 88b, as shown in FIG. 2, the reduction tube and cam 88 arerotatively locked together.

Referring still to FIGS. 2 and 4, rotatively mounted on shaft 74 beyondreduction tube 110 is a bushing 162 (not shown in FIG. 4) and encirclingthe bushing adjacent the reduction tube flange is a circular ring orplate 164 which is more or less a mirror image of ring 132. As such ithas a circular bearing race 166 for receiving the ball bearings 114carried by the tube 110. The tubes 110 and 124 and the rings or plates116 and 132 comprise torque take-off means with tube 110 thereofproviding the rotary output to the strain wave generator 86, while thebushing 162 and ring 164 actually forming part of the clutch section 28to be described later.

A generally rectangular key 168 is provided to rotatively couple therings or plates 132 and 164. One end of the key is arranged to fit in akeyway 172 in the rim of ring 132, while the opposite end of the keyseats in a similar keyway 174 formed in the rim of ring 164 and projectsan appreciable distance beyond that ring. A relatively long notch 175 isformed in the underside of the key to provide clearance for tubes 110and 124 and the intervening ring 116.

As best seen in FIGS. 2 and 8, encircling the key is a sleeve 176 (notshown in FIG. 4) having an internal keyway 177 for receiving key 168.The sleeve is positively secured to the key by means of set screws 178extending through countersunk openings in the sleeve and turned downinto threaded openings 184 in the top of the key. As best seen in FIG.8, the end 176a of sleeve 176 outboard of bearing 136 and facing disk138 is beveled and provided with teeth, thereby forming a female bevelgear 188 located directly opposite the comparable gear 104 in collar102.

Referring now to FIGS. 2, 4, 7 and 8, the speed reduction disk 138 inthe planetary section 26 is actually composed of two shell halves 138aand 138b which are more or less mirror images of one another so thatwhen the two halves are brought together and secured at their edges,they define an annular space 192 inside the disk. Also each disk half isformed with a generally rectangular slot 194 near its edge, the twoslots being in register thereby forming a window through the disk.

Positioned between the two slots 194 (FIGS. 4 and 7) inside the disk isa so-called gear shift key 196. The key is pivotally mounted to the diskby means of a pin 198 extending through an opening in the outboard endof the key with the ends of the pin being staked in registering passages202 in the two disk halves near their rims. The pin 202 is oriented sothat it is perpendicular to the sidewalls of the slots 194. This permitsthe key to pivot between one position wherein the key projects outthrough the slot 194 at one side of the disk as shown in FIG. 2 andanother position wherein it projects out through the slot 194 at theopposite side of the disk, as seen in FIG. 8. The opposite sides of thekey which project through the slots are formed with sets of male bevelgear teeth 206 and 208. When the key is in its said one position, itsteeth 206 mesh with the bevel gear 104 in collar 102. When the key is inits other position, its teeth 208 mesh with the bevel gear 188 in sleeve176.

Positioned in the annular space 192 inside the disk is a tightly coiledlength of spring wire whose ends overlap to form a circular spring 210.One side of the spring seats in a notch 212 formed in the end of thegear shift key opposite its pin 202, as best seen in FIGS. 7 and 8. Adiametrically opposite point on the spring is captured by lateralprojections 214 in the two disk halves. The positions of the notch 212and the pin 202 are such that the gear shift key is over center whenmoved near either of its two aforesaid positions. In other words, thekey is a bistable element whose teeth either mesh with bevel gear 104 incollar 102 or with bevel gear 188 in sleeve 176.

The gear shift key is switched between its two stable positions by meansof a channel-shaped cam 216 shown in detail in FIG. 9. The cam ispivotally supported in a generally U-shaped channel 218 by means of apivot pin 222 extending through openings 224 in the opposite side wallsof the cam and through registering openings 226 in the opposite sidewalls of the channel. The channel 218 is spaced in front of clutchsection 28 by spacers 227 and supported therefrom by bolts 229 (FIG. 2)so that cam is located directly below the gear shift key 196.

As best seen in FIG. 9, the opposite side walls of the cam are formedwith two raised lobes 232 and 234 which are symmetrically disposed aboutthe pivot pin 224. The shapes of the lobes are such that when the cam ispivoted in one direction about its pin 222, its lobe 234 is raised andengages the gear shift key 196 and moves the key over center so that itsnaps into gear 104 as shown in FIGS. 2 and 7. On the other hand, whenthe cam is tilted in the opposite direction, its lobe 232 is raised andengages the opposite side of the key, thereby snapping the key to itsother position wherein it meshes with sleeve 176 as shown in FIG. 8. Thecam is tilted between its two positions by the apparatus control section30 as will be described later.

It should also be appreciated at this point that during operation of theapparatus, the disk 138 may rotate with tube 124 relative to cam 216 andchannel 218. Therefore the key 196 can only be switched between its twopositions as it swings by the cam 216. If the key is already in thedesired position as selected by the control section 30 when it swingspast the cam, the cam will have no effect on the key. However, if thekey is in its opposite position, as soon as it strikes the cam, it willbe switched to its desired position.

As mentioned at the outset, the function of the planetary section 26 isto enable the clutch section 28 to rotate the wave generator 88 relativeto the deformable gear 62 in drive section 24 at two different speeds asselected by control section 30. This is accomplished by means of theshift key 196 through the action of the reduction tube 110, the drivetube 124 and their associated bearings and races. These are not simplebearing systems as appears in the drawings. Rather, the races on rings116, 132 and 164 may have slightly different radii. Moreover, thebearings 114 on tube 110 are lubricated by or coated with a speciallubricant which, when subjected to high pressure, has a highco-efficient of traction. This occurs when the bearings are subjected toan axial load as is the case in the present apparatus. Under relativelylow pressure, the material functions as an ordinary lubricant. Asuitable lubrication material to produce this effect is sold by MonsantoChemical Co. under the trade name SANTOTRAK. With an axial thrust on thecoated bearings of about 40 pounds, this material produces about 4pounds of rotary or shear force. As such, the bearings and their racesfunction more or less like meshing gear teeth in a planetary gear sytem.

More particularly and referring to FIGS. 2 and 4, assume shaft 74 isstationary. When ring 164 in clutch section 28 rotates relative to theshaft, so does ring 132 and sleeve 176 because of the interconnectingkey 168. If the shift key 196 is in its FIG. 8 position wherein itmeshes with gear 188 on the sleeve 176, the disk 138 also rotates alongwith those other components. As just described, the bearings 114 and 128on tubes 110 and 124 respectively and the opposing bearing races onrings 116, 132 and 164 function effectively as gear teeth. With this inmind, it is apparent that reduction tube 110 will rotate with ring 164by virtue of its bearings 14 engaging that ring. Also ring 116 capturedtractively between bearings 114 and 128 will rotate along with thoseother components. Since all of these components are rotating in the samedirection and speed, the wave generator 86 coupled to reduction tube 110will rotate along with them in the same direction and at the same speedas ring 164. Thus there is a one-to-one speed coupling between ring 164and the wave generator.

Assume now that the key 196 is switched to mesh with gear 104 on collar102 attached to input shaft 34 as shown in FIGS. 2 and 7. Rings 132 and164 and sleeve 176 rotate together as before. Now, however, disk 138 aswell as drive tube 124 coupled thereto are rotatively locked together.Further the outer ring gear 68 and the input shaft 34 coupled theretorotate at 79/80 of the shaft speed due to the 80:1 speed reductionprovided by drive section 24. Therefore, for purposes of this discussionthey can be considered to rotate essentially at the shaft speed, i.e.zero in this example. Therefore the drive tube 124 is also essentiallystationary.

With this arrangement, the rotating ring 132 acting through thetractively captured bearings 128 in tube 124 cause ring 116, alsoengaged by bearings 128, to rotate at the same speed but in the oppositedirection. As such, it also rotates oppositely to ring 164 keyed to ring132. This means that the bearings 114 on reduction tube 110 aretractively captured on each side by rings 116 and 164 which rotate atthe same speed but in opposite directions.

If the radii of the races on rings 116, 132 and 164 were equal, then thebearings 114 would simply rotate about their axes and no torque would becoupled to tube 110 and wave generator 86 would be stationary along withshaft 74. However, as alluded to previously, those radii are not equal.Rather as shown in the FIG. 10 diagram, the balls 128 on tube 124contact races 134 and 122 on rings 132 and 116 respectively at pointsdisplaced from the ball axis by distances -A and +A respectively, theball axis being spaced a unit distance from the shaft 74 axis. In otherwords, race 122 has a slightly larger radius than race 134. On the otherhand, the balls 114 on tube 110 contact races 118 and 166 on rings 116and 164 respectively at points displaced from the ball axis by distances+B and -B respectively, distances A and B being different.

Using standard equations for planetary gearing, it can be shown that,when the keyed-together rings 132 and 164 are rotated relative to shaft74, the rotation of the tube 110 with respect to those rings may beexpressed as follows: ##EQU1##

Therefore by inserting various values of A and B simply by substitutingdifferent rings, the planetary section 26 can be made to producedifferent speed reduction ratios for different applications of theapparatus 10. The races on the two rings can even be angled in the samedirection as specifically shown by the dotted line races 134' and 122'in FIG. 10. The following Table I shows various ratios obtained byplanetary section 26 for different values of A and B.

                  TABLE I                                                         ______________________________________                                                                      Speed                                                                         Reduction                                       Ex.     A(inch)       B(inch) Ratio                                           ______________________________________                                        1       .027          .030    357:1                                           2       .028          .030    526:1                                           3       0             030      33:1                                           4       0             .015     67:1                                           5       .025          .030    208:1                                           6       0             .050     20:1                                           ______________________________________                                    

Planetary section 26 in the illustrated apparatus 10 produces areduction ratio of 20:1. Therefore, the race radius values of ExampleNo. 6 may be employed. Section 26 will produce that speed reductionwhichever direction the rings 132 and 164 rotate relative to the shaft74. Therefore, when the shift key 196 is in its FIG. 2 position so thatsection 26 is operative, the total speed reduction ratio between thewave generator and the output shaft is the product of that provided byplanetary section 26, i.e. 20:1 and that provided by drive section 24,i.e. 80:1 or a total of 1600:1 in this example.

Referring now to FIGS. 2, 4 and 5, the clutch section 28 includes acircular annular plate 232 whose lower edge is bent back to form ahorizontal tab 232a which rests on the apparatus control section 30.Also formed at the inner edge of the plate 232 are three laterallyextending tabs 234 spaced 120° apart around the plate. These tabs areslightly curved (FIG. 5) having a radius corresponding to the radius ofthe inner edge of the plate. Also each tab 234 has a generallyrectangular slot 236 extending in from one end. Each slot cages a dowelor roller bearing 237. During assembly, the dowels may be temporarilyretained in their slots by an application of grease.

Plate 232 loosely engages over sleeve 176 of the planetary section 26.Positioned just beyond plate 232 is a flat annular clutch dog 238 whichloosely encircles the tabs 236 in the frame plate. To assure adequateclearance between those elements, the clutch dog 238 is formed withthree notched sectors 242 (FIG. 4) at its inner edge spacedapproximately 180° apart and located directly adjacent the tabs 234. Theclutch dog is also formed with a depending handle 238a by which the dogcan be cocked or rotated. Also directly above handle 238a a key 238bprojects radially inward. This key is arranged to fit in a keyway 244 ina clutch ring 246 positioned just beyond the dog so that the dog and theclutch ring 246 move together.

As best shown in FIS. 2, 4 and 5, the outside surface of the clutch ring246 is formed with three flats 248 which are spaced 120° apart aroundthe ring and are located directly opposite the tabs 234 on the frameplate 232. Also, the clutch ring has a radial flange 246a at its far endwhich extends radially outward beyond the flats. The ring 246 isarranged to nest in a clutch plate 252 whose face is formed with anannular channel 254 which is dimensioned to relatively loosely receivethe clutch ring 246. The clutch plate is formed with a central opening256 (FIG. 5) which is arranged to engage over the larger end of thebushing 162 on shaft 74.

With the clutch ring 246 seated in the clutch plate 252, the clutchplate is engaged over bushing 162 beyond the frame plate 232 and dog 238as shown FIGS. 2 and 5 so that the dowels 237 are captured between theouter wall of groove 254 in the clutch plate and the flange 246a andflats 248 on the clutch ring as best seen in FIG. 2. Also as shown inFIGS. 4 and 5, the inner wall of the clutch plate groove 254 is formedwith an axial keyway 262 which is arranged to receive the projecting endof the key 168 extending out from the planetary section 26. Thus, whenthe clutch plate 252 is seated, it is rotatively locked to the ring 164,ring 132 and sleeve 176 by the key 168. On the other hand, the dog 238and clutch ring 246 are rotatively locked together, but free to cockrelative to the aforesaid components.

Positioned on the output shaft 74 beyond bushing 162 is a drive tube 268having a radial flange 268a formed with a circular array of openings 272which cage an array of ball bearings 274. These bearings are lubricatedwith the same tractive material described above. Tube 268 is rotativelykeyed to shaft 74 at 275 and thus rotates with the shaft, but it maymove longitudinally to some extent relative thereto.

As shown in FIG. 2, the bearings 274 carried by tube 268 ride in acircular race 276 in the adjacent face of the clutch plate 252. Also,encircling the rim of the cage is an annular so-called reference plate277 carrying a brass bushing 278 at its inner edge which snugly engagesaround the tube flange 268a, but permits its rotation.

Positioned beyond and encircling the neck portion of tube 268 is asecond clutch plate 282 which is more or less a mirror image of plate252 in that it is formed with a circular race 284 in its face adjacenttube flange 268a which functions as a race for the bearings 274. Also ithas a circular groove 286 formed in its opposite face more or less inline with the groove 254 in plate 252.

Seated in groove 286 is a clutch ring 288 which is more or less a mirrorimage of the clutch ring 244 in that its outer surface is formed withthree flats (not shown) spaced 120° apart which are aligned with thesimilar flats 248 in ring 246. Positioned beyond the clutch ring 288 isan annular dog 294 more or less identical to dog 238 in that it has adepending handle 294a and an upstanding key 294b which engages in akeyway (not shown) formed in the rim of the clutch ring 288 at thebottom of the ring. Thus the dog and clutch ring rotate together aboutthe shaft 74 axis.

Positioned beyond the clutch ring is an end plate 302 which isessentially a mirror image of plate 232 in that its lower edge is bentlaterally to form a tab 302a that seats on the control section 30. Theplate is provided at its inner edge with three laterally extending tabs304 spaced 120° apart located directly opposite the flats on the clutchring 288. As with plate 232, these tabs cage a set of three pins ordowels 308 which extend between the clutch ring flats and the outer wallof groove 286 in the clutch plate 282.

The three plates in section 28 are held together at their tops by a pinor bolt 305 extending through aligned openings in those plates, suitablespacers 307 being provided between the plates to permit rotation of theclutch plates 252 and 282. The bottoms of those plates are held togetherby the bolts 229 supporting channel 218, appropriate spacers (not shown)being provided there as well.

Clutch section 28 operates as follows. The drive tube 268 rotating withshaft 74 normally causes the adjacent clutch plates 252 and 282 torotate in the same direction and speed. This is due to the tractivenature of the lubricant on bearings 274 when the system is under axialload as described above. Assume, for example, that tube 268 is rotatingclockwise at speed V. If now the clutch plate 252 is stopped, the clutchplate, as well as the rings 132 and 164 in apparatus section 26 keyedthereto, will have a velocity of -V relative to shaft 74.

On the other hand, if plate 252 is allowed to rotate and plate 282 isstopped, then plate 252 and the rings 132 and 164 will rotate relativeto the shaft 74 at a speed equal to the surface speed of the bearings274, i.e. +2V minus the speed of tube 268, i.e. -V, or +V. Thus bystopping one or the other of the clutch plates, the ring 164 that drivesthe planetary section 26 can be made to rotate at a speed V in onedirection or the other relative to the output shaft.

The clutch plates 252 and 282 are stopped by actuation of the clutchdogs 238 and 294, respectively by control section 30. Since the two dogsoperate identically, we will only describe the functioning of the formerone in detail.

Referring to FIGS. 5 and 6, when the dog 238 is in its disengagedposition shown in FIG. 5, the flats 248 of clutch ring 248 are centeredon the dowels 237 caged by the stationary tabs 234. In this position,the dowels 237 do not bear appreciably against the wall of the groove254 in the clutch plate. Therefore the plate is free to rotate withdrive tube 268 as described above. Assume for example, it is rotatingclockwise as shown by the arrow in FIG. 5.

If now the dog 238 is rotated slightly to its engaged positionillustrated in FIG. 6, the clutch ring 246 is cocked so that its flats248 wedge the dowels 237 against the groove wall of clutch plate 252.This wedging action brings the plate 252 to a rapid but smooth stop. Ina general sense, then, the clutch ring, clutch plate and dowelscooperate to form a sprag clutch or brake. In a conventional spragclutch, however, when the clutch is being engaged the rings or plates onthe opposite sides of the pins or dowels move in the same direction sothat the pins tend to be rolled and squeezed into the wedge. This makesit relatively difficult to disengage the usual sprag clutch. Here,however, when engaging the clutch, the plate 252 and ring 246 arerotating in opposite directions so that the dowels 237 have lesstendency to become "bound up" in the wedge between those two members.Accordingly, it is easier to disengage the clutch by returning the dogto its original FIG. 5 position. Since the clutch is moved to itsengaged position by a solenoid in control section 30 and disengaged by areturn spring acting on the dog, a lighter spring may be used permittinguse of a smaller and less expensive solenoid than could be used tooperate a conventional sprag clutch.

Clutch dog 294 is operated in the same way to wedge the dowels 308between clutch ring 288 and the groove wall in clutch plate 282 to stopthat plate and that clutch arrangement has the same benefits as the onedescribed above.

Referring again to FIG. 2, encircling a reduced diameter neck portion268b of the drive tube 268 is a thrust bearing 309 and beyond the thrustbearing, a stack of Belleville washers 310 encircle shaft 74. Finally,beyond the washers is a threaded nut 312 which is turned down ontocorresponding threads 313 at the end of shaft 74. If desired, a greasefitting 316 may be screwed onto the end of the shaft and the shaft maybe formed with an axial passage (not shown) and connecting lateral ports(not shown) in order to deliver the tractive-type lubrication to thebearings on the various rotary tubes. Desirably also, the nut 312 ispositively locked to the shaft by a set screw 318 to firmly fix itsposition thereon.

As the nut 312 is turned down onto the shaft, the Belleville washers 310apply an axial load to the wave generator 86 by way of the variousrotary components slidably mounted on the shaft 74 in sections 26 and28. The wave generator 88 and the bearing race 98 carried thereby arethus biased against the deformable inner drive gear 62 therebydeflecting section 62c of the gear into its elliptical shape so that itsteeth engage the teeth of the rigid outer ring gear 68 at diametricallyopposite locations thereon as described above. Because of the specialshape of gear 62 described above, an axial force of only about 40 poundssuffices for this purpose. That force is also large enough to createsufficient pressure at the points of contact between the bearings andtheir races to impart the required coefficient of traction to thebearing lubricant at those locations. However the axial load is lowenough so that there is minimum wear of the axially loaded bearings,rings and washers and minimum energy loss when rotating those elementsrelative to one another.

The applied axial load can also be adjusted to change the differencebetween the major and minor axes of the deformable gear 62 in order toadjust the angulation factor thereof discussed above. In other words, ifthe axial force is increased, the difference between the major and minoraxes in a corresponding radial plane also increases in proportion to thegear spring constant so that the spacing between the unengaged toothpoints is greater. Conversely, if the axial load is decreased, thespacing between those tooth points is decreased.

It should be emphasized at this point also that the components of thepresent apparatus are not critically dimensionally dependent since theyare force loaded. Accordingly, there is no tolerance accumulationproblem. In other words, the interactive components of the apparatus,including the wave generator 86, the ring gears and bearing elements,clutch plates, etc. are all face-to-face-contacting components which aremaintained in the proper engagement by the axial load imposed by theBelleville washers 310. The apparatus does not utilize any rim-engaginggears, spiders or other dimensionally critical gear trains found in theprior rotary shaft control apparatus. Consequently apparatus 10 is mucheasier and less expensive to manufacture and assemble than those priordevices. Furthermore for the same reasons, the apparatus suffers littleor no backlash even if the drive section 24 gear teeth or the bevelgears in section 26 become worn in use because the gears are alwaysaxially biased together.

Referring to FIGS. 1, 2 and 7, the control section 30 comprises ahousing 340 which is secured to the plate flanges 232a and 302a by bolts344 (FIG. 2), suitable slots being provided in the top wall of thehousing to accommodate the dog legs 238a and 294a. The housing 340contains three solenoids 352, 354 (FIG. 2) and 356 which control theabove-described movements of the dogs 294 and 238 in section 28 and thegear shift cam 216 in section 26 respectively. Those three solenoidsare, in turn, energized by the switch buttons 54, 53 and 56 respectivelydepicted in FIG. 1. The circuitry for energizing the solenoids by meansof the switches and the mechanical linkages between the solenoidarmatures and the dogs and cam are conventional and will not be detailedhere. Suffice it to say that these buttons are normally biased upwardsin their solenoid-off positions so that dogs 238 and 294 are in theirdesengaged positions shown in FIG. 5 and cam 216 is in its key-engagingposition of FIG. 2.

During normal operation of the apparatus, the buttons 53, 54 and 56 arein their raised positions so that the dogs 238 and 294 are positioned topermit rotation of the two clutch plates 252 and 282. Also the shift cam216 is positioned to place the shift key in engagement with the bevelgear 104 as shown in FIG. 2. We will assume that the shaft 74 and roller12 are rotating clockwise at a speed V and that shaft speed will be usedas the speed reference for purposes of the following discussion. In thisnormal mode of operation, as described above, the rotary components ofthe clutch section, namely clutch plates 252 and 282 rotate at the shaftspeed V. That motion is coupled by key 168 to the rotary components ofthe planetary section 26 which also rotates at the shaft speed V.Likewise the reduction disk 138 coupled to the bevel gear 104 by theshift key 196 also rotate at the same shaft speed, as does the inputshaft 34 to which the that bevel gear is connected.

With all of the rotary components of the apparatus turning together,there is no relative motion between the wave generator and the inner andouter ring gears in the drive section 24. Accordingly, the input shaft34 is essentially keyed to the output shafts 32 and 74 and the rollershaft 14 coupled thereto. Thus torque is coupled directly between theinput and output shafts by way of the ring gears 62 and 68. No loadtorque is coupled to the wave generator 86 or through the wave generator86 to the apparatus sections 26 and 28. Moreover, in being stationaryrelative to one another, the rotary components of those sections do notsuffer any appreciable amount of wear during normal operation of theapparatus.

Assume now that it is desired to bring the arrow A on roller 12 in FIG.1 into registration with the arrow C on sprocket 42. This involvesadvancing the sprocket relative to the roller. Assume further that it isdesired to make that correction at the faster rate of speed, at leastuntil the point of registration is at hand. To accomplish this, theoperator first presses the SPEED CONTROL button 56. This energizes thesolenoid 356 (FIG. 7) which switches cam 216 from its normal positionillustrated in FIG. 2 to the position illustrated in FIG. 8 wherein itslobe 232 is in the raised position.

As soon as the rotating disk 138 rotates the shift key 196 around to itslowermost position shown in FIG. 8, the key will engage the cam lobe 232and be switched to the position shown in FIG. 8 wherein it meshes withthe bevel gear 188 on sleeve 176. As long as the fast speed correctionmode is desired, the speed control button 56 must be maintained in itsdepressed position. Otherwise it will automatically return to its raisedposition and cause the cam 216 to return the shift key 198 to itsoriginal position illustrated in FIG. 2 during the next revolution ofthe disk 138.

Since the arrow A leads arrow C, the operator next depresses the RETARDbutton 54. As described previously, that energizes solenoid 352 whichshifts the dog 294 to a position similar to the one shown in FIG. 6 withthe result that the clutch plate 282 is brought to a stop. As describedabove, this causes the opposite clutch plate 252 to rotate relative toshaft 74 at a speed of +V where V is the speed of shaft 74. Accordinglywith the shift key 196 in that position, that rotation is coupleddirectly through planetary section 26, which imparts no speed reductionto the wave generator 86 which therefore rotates at a speed +V relativeto the shaft 74. This rotation of the wave generator causes the inputshaft 34 and its sprocket 42 to be advanced in phase relative to theshafts 32, 74 and 14 at the reduced speed ratio characteristic of theparticular strain wave drive 24, i.e. 80:1 in this example. That isequivalent to about 900° per minute when shaft 74 is running at 200 rpm.

If the operator greatly overshoots the point of registration, whilemaintaining button 56 depressed, he can release the RETARD button 54disengaging its clutch assembly and depress the ADVANCE button 53. Thatwill actuate solenoid 354 and move dog 238 to its engaged positionillustrated in FIG. 6 thereby stopping the clutch plate 252 so that itsspeed relative the shaft 74 is -V. That relative motion is coupleddirectly through planetary section 26 to wave generator 86 in exactlythe same way as described above and will cause the input shaft 34 andsprocket 42 to be retarded in phase relative to roller 12 at 1/80th ofthe shaft speed V.

Assume now that the operator wishes to bring the roller arrow A intoregistration with the sprocket arrow C before any overshoot at theslower rate to achieve precise registration. He releases the speedcontrol button 56 causing solenoid 356 to return cam 216 to its normalposition illustrated in FIGS. 2 and 7. In that position, its lobe 234engages the shift key 196 upon the next revolution of the disk 138thereby switching the key to its position shown in those figures so thatit meshes with the teeth 104 and collar 102. The operator now depressesthe RETARD button 54 thereby stopping the clutch plate 282 as describedabove. This causes the clutch plate 252 to rotate relative to the shaft74 at a speed of +V. However, with the shift key 196 in that position,as described above the planetary section 26 effects a 20:1 speedreduction so that the strain wave generator 86 is rotated relative toshaft 74 at a speed of 1/20V. When multiplied by the 80:1 speedreduction characteristic of the drive section 24, this causes the inputshaft 34 and the sprocket 42 to be advanced in phase relative to roller12 at a speed of 1/1600V. Consequently, the sprocket arrow A will bemoved extremely slowly into registration with the sprocket arrow C sothat the precise point of registration can be achieved without anyovershoot. At that point, the operator releases the button 54.

In the same fashion, the apparatus can be controlled to advance theroller relative to the sprocket at the slower rate by depressing button53 with button 56 in its raised position. The operation of the apparatusduring that mode is exactly as described above except that clutch plate252 is stopped causing the strain wave generator 86 to rotate in theopposite direction at 1/20V thereby causing the shaft 34 and sprocket 42to rotate in the opposite direction from the roller at a relative speedof 1/1600V.

It is important to note that the rotary clutch plates 252 and 282 haveidentical rotary moments of inertia. Therefore, their angular momentumsare identical. When one plate is stopped to effect a phase correction,its angular momentum is transferred to the other plate which proceeds torotate twice as fast so that the total angular momentum of the rotarysystem as a whole remains the same. Thus virtually no shock istransmitted to the machine frame when switching between the ADVANCE andRETARD modes of operation. Furthermore, the clutch section of theapparatus is fail safe in that even if the operator inadvertentlydepresses the ADVANCE and RETARD buttons at the same time, therebylocking both clutch plates 252 and 282, the bearings 274 simply skid inthe clutch plate races.

Likewise, the switching of the apparatus between its two speed modesresults in no appreciable amount of shock to the machine frame. This isbecause at the time of switching, the disk 138 and the bevel gears 102and 188 are rotating at essentially the same speed.

It will be seen from the foregoing, then, that the apparatus 10 permitsthe phase and speed of a rotary shaft to be very closely controlledrelative to the position of another rotary member. The apparatus ismounted directly on the shaft which it controls and effects its speedand phase corrections without requiring a secondary power source.Moreover, the apparatus is compact so that it can be retrofit onexisting machinery without requiring relocation thereof. Additionallythrough its novel clutch and planetary mechanisms, the apparatus permitsspeed and phase corrections at a plurality of speeds. Further, becauseof the special hour-glass shape of gear 62, the components of theapparatus suffer minimum stress and wear. Yet with all of theseadvantages, the apparatus as a whole and drive section 24 in particularare relatively inexpensive to make and assemble. Furthermore, for thereasons described above, they require a minimum amount of maintenanceand repair.

It will thus be seen that the objects set forth above, among those madeapparent from the preceding description are efficiently attained. Also,certain changes may be made in the above construction without departingfrom the scope of the invention. For example, instead of using manualbutton switches in the control section 30, the solenoids in the controlsection can be actuated from a remote location by suitable limit sensingmeans or other sensing devices to cause apparatus 10 to control theshaft 14 relative to a reference in response to various conditions suchas web tension changes, the height of a liquid in a process tank, etc.Also if the high gear reduction provided by section 26 is not needed ina given application, the section 26 including collar 102 on the inputshaft can be replaced by an annular adaptor one side of whichcorresponds to the notched end of tube 110 and the other side of whichis provided with a lateral key like the projecting end of key 168 sothat the adapter rotatively couples the wave generator to ring 252 inclutch section 28. Of course spacer rings would have to be providedbetween thrust bearing 309 and the Belleville washers 310 to axiallyload the wave generator. Therefore, it is intended that all mattercontained in the above description or shown in the accompanying drawingsshall be interpreted as illustrative and not in a limiting sense.

It is also to be understood that the following claims are intended tocover all of the specific features of the invention herein described.

What is claimed as new and desired to be secured by Letters Patent ofthe United States is:
 1. Torque coupling apparatus for use particularlyfor rotary shaft control comprisingA. a rotary shaft for rotating in agiven direction, B. a rotary bearing unit rotatively fixed to said shaftand having an array of bearing elements projecting out from both facesof the bearing unit, C. a first plate rotatively mounted on said shaftcoaxially with respect to the bearing unit, said plate being formed witha race in one face thereof engaging the bearing elements projecting fromone face of the bearing unit, D. a second plate rotatively mounted onsaid shaft coaxialy with respect to the bearing unit, said second platebeing formed with a second race in one face thereof engaging the bearingelements projecting out from the other face of the bearing unit, E.torque take-off means rotatively mounted on said shaft and coupled tothe second plate, said torque take-off means including rotary outputmeans, F. a tractive-type lubricant applied to the bearing elements,said lubricant exhibiting significant shear strength when subjected to aselected pressure, G. means on said shaft for axially loading the platesagainst the bearing unit so as to exert said selected pressure so thattorque may be coupled between the bearing unit and the plates, H. meansfor slowing or stopping the rotation of said first plate when said shaftis rotating in said given direction whereby the output means areadvanced in phase relative to said shaft, and I. means for slowing orstopping the rotation of the second plate when said shaft is rotating insaid given direction whereby said output means are retarded in phaserelative to said shaft.
 2. Torque coupling apparatus comprisingA. arotary shaft for rotating in a given direction, B. a rotary bearing unitrotatively fixed to said shaft and having an array of bearing elementsprojecting out from opposite faces thereof; C. a first plate rotativelymounted on said shaft, said plate being formed with a race in one facethereof engaging the bearing elements projecting from one face of thebearing unit; D. a second plate rotatively mounted on said shaft, saidsecond plate being formed with a second race in one face thereofengaging the bearing elements projecting from the opposite face of thebearing unit; E. a tractive-type lubricant applied to the bearingelements, said lubricant exhibiting significant shear strength whensubjected to a selected pressure; F. means on said shaft for axiallyloading said plates against the bearing unit in order to exert saidselected pressure so that torque may be coupled between the bearing unitand the plates; G. torque take-off means including rotary output meansrotatively mounted on said shaft for coupling torque from one of saidplates to a load so that, when said shaft is rotated in said givendirection and the angular velocity of said output means change, angularmomentum is transferred to the other plate so as to reduce the shockload on the system due to said angular velocity change.
 3. The systemdefined in claim 2 and further including means for braking said oneplate and said output means whereby, due to said angular momentumtransfer, minimal torque is imparted by the brake to the rotary system.4. The apparatus defined in claim 3 wherein said braking meanscompriseA. stationary reference means, and B. a clutch acting betweensaid one plate and the reference means so as to rotatively coupletogether said one plate and said reference means.
 5. Torque couplingapparatus for use particularly for rotary shaft control comprisingA.first and second rotary plates mounted to rotate about a common axis, B.means defining a circular wall in one face of each of said plates, C.first and second rings mounted for rotation about said axis, said ringseach having peripheral flats disposed opposite said wall in a differentone of said plates, first and second sets of pins, the number of pins insaid sets corresponding to the number of flats on said rings, E. firstand second pin support means mounted for rotation relative to saidplates, said first support means supporting the first set of pins forrotation about their axes between the flats of said first ring and saidwall of said first plate, said second support means supporting thesecond set of pins for rotation about their axes between the flats ofsaid second ring and said wall of said second plate, F. means forrotating each said ring in a direction opposite to the rotation of itscorresponding plate whereby the corresponding pins are wedged betweenthe end portions of the corresponding flats and said wall of thecorresponding plate so as to rotatively couple the plate andcorresponding support means and whereby when each said ring is rotatedin the opposite direction from its corresponding plate, said plate andits corresponding support means are decoupled by a rolling action ofsaid corresponding pins, G. bearing means rotatively positioned betweensaid plates, said bearing means including bearing elements projectingout from opposite faces of said bearing means toward said plates, H.circular races formed in the faces of said plates for receiving saidbearing elements, I. a tractive lubricant coating said bearing elements,said lubricant exhibiting a significant amount of shear strength whensubjected to pressure, and J. means for axially loading the plates andbearing means so as to exert said pressure on the lubricant whereby whenthe angular velocity of one plate is changed, that velocity change iscoupled to the other plate.
 6. The apparatus defined in claim 5 whereinthe angular moments of inertia of the two plates are substantially thesame.
 7. The apparatus defined in claim 5 wherein the angular moments ofinertia of the two plates are different.
 8. Torque coupling apparatusfor use particularly for rotary shaft control comprisingA. a rotarybearing unit having an array of bearing elements projecting out fromboth faces of the bearing unit, B. a shaft rotatively coupled to thebearing unit, C. a first plate rotatively mounted concentrically on theshaft coaxially with respect to the bearing unit, said plate beingformed with a race in one face thereof engaging the bearing elementsprojecting from one face of the bearing unit, D. a second platerotatively mounted concentrically on the shaft coaxially with respect tothe bearing unit, said second plate being formed with a second race inone face thereof engaging the bearing elements projecting out from theother face of the bearing unit, E. torque take-off means rotativelymounted concentrically on said shaft, said torque take-off means beingcoupled to the second plate and including rotary output means, F. atractive-type lubricant applied to the bearing elements, said lubricantexhibiting significant shear strength when subjected to a selectedpressure, G. means for axially loading the plates against the bearingunit so as to exert said selected pressure so that torque may be coupledbetween the bearing unit and the plates, H. means for slowing orstopping the rotation of said first plate when the bearing unit isrotating in a given direction about its rotary axis whereby the outputmeans rotate in one direction relative to the bearing unit, and I. meansfor slowing or stopping the rotation of the second plate when thebearing unit is rotating in said given direction so that the outputmeans rotate in the opposite direction relative to the bearing unit. 9.Torque coupling apparatus for use particularly for rotary shaft controlcomprisingA. a rotary bearing unit having an array of bearing elementsprojecting out from both faces of the bearing unit, B. a first platerotatively mounted coaxially with respect to the bearing unit, saidplate being formed with a race in one face thereof engaging the bearingelements projecting from one face of the bearing unit, C. a second platerotatively mounted coaxially with respect to the bearing unit, saidsecond plate being formed with a second race in one face thereofengaging the bearing elements projecting out from the other face of thebearing unit, D. torque take-off means coupled to the second plate andincluding rotary output means, E. a tractive-type lubricant applied tothe bearing elements, said lubricant exhibiting significant shearstrength when subjected to a selected pressure, F. means for axiallyloading the plates against the bearing unit so as to exert said selectedpressure so that torque may be coupled between the bearing unit and theplates, G. means for slowing or stopping the rotation of said firstplate when the bearing unit is rotating in a given direction about itsrotary axis whereby the output means rotate in one direction relative tothe bearing unit, and H. means for slowing or stopping the rotation ofthe second plate when the bearing unit is rotating in said givendirection so that the output means rotate in the opposite directionrelative to the bearing unit, said slowing or stopping meansincluding(1) stationary reference means, and (2) clutch means actingbetween one of said plates and the reference means so as to rotativelycouple said one plate and reference means together.
 10. The apparatusdefined in claim 9 whereinA. each plate is formed with a circular wallin its other face, and B. each clutch means comprise(1) a ringrotatively mounted relative to a plate, said ring having one or moreperipheral flats disposed opposite said plate wall, (2) a correspondingnumber of pins, (3) means secured to the reference means for rotativelysupporting each pin about its axis between a said flat and said platewall, there being sufficient clearance between said pins and said platewall to permit rotation of said plate, and (4) means for rotating saidring relative to said reference means so as to wedge said pins betweensaid flats and said plate wall thereby rotatively coupling the plate tothe reference means.